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Full text of "Handbook Of Chemical Engineering - I"

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THE TRANSPORTATION OF GASES                        161
become less as the volume represented by that point increases. Assume as a basis the delivery of a fan 1 ft. in diameter at 51.5 per cent mechanical efficiency where it delivers 1,500 cu. ft. per minute against 1.67 in. static pressure at 1,274 r.p.m., requiring 0.95 b.hp. Operating at the same point on the mechanical efficiency curve but against 1.5 in. pressure would require the speed to be lowered in proportion to the square root of the decrease in pressure, or be multiplied by the factor 0.947. The volume would decrease in the same proportion and the brake horsepower as the third power of 0.947, giving a delivery of 1,412 cu. ft. per minute against 1.5 in. static pressure at 1,200 r.p.m. and requiring 0.81 b.hp. Since a delivery of 50,000 cu. ft. per minute is desired at this static pressure of 1.5 in., the peripheral speed will remain the same but the diameter will vary as the square root of the increase in volume. The speed in revolutions per minute will vary inversely as the diameter and the brake horsepower directly as the volume, giving a fan 5.95 ft. in diameter to deliver 50,000 cu. ft. per minute against 1.5 in. at 202 r.p.m. and requiring 28.7 b.hp. Since the diameter obtained is an odd figure, it might be decided to use a fan of even diameter, say 6 ft., in which case its performance would be calculated by considering that a similar fan 1 ft. in diameter at the same mechanical efficiency would deliver a volume in proportion to the square of the diameters, or Jie of 50,000, which is 1,390 cu. ft. per minute, against 1.5 in. pressure. Since the area of outlet is 0.59 sq. ft. the velocity in the outlet is 2,350 ft. per minute and the velocity pressure 0.347 in., hence the ratio of static to impact pressure is then 0.813. Figure 12 contains a curve of this ratio, which facilitates such computations and does away with cut-and-try methods. From Fig. 12 it is seen that when the ratio of static to impact pressure equals 0.813, the l-ft.-diameter fan delivers 1,465 cu. ft. per minute against 1.69 in. static pressure for 0.93 b.hp. at 1,274 r.p.m. Now to deliver 1,390 cu. ft. per minute against 1.5 in., the fan speed would drop in direct proportion to the volume, or to 1,209 r.p.m., and the brake horsepower as the cube, or to 0.79 b.hp. From this the speed and horsepower of the 6-ft. fan when delivering 50,000 cu. ft. per minute against 1.5 in. are found to be 201 r.p.m. and 28.5 b.hp.
Figure 13 covers a double-inlet fan with wheels and casing of the same proportions as the fan of Fig. 12 the three wheels tested in this casing had respectively 60, 48 and 36 blades. The one with 60 blades was the same wheel as used in Fig. 12. The effect of making the casing with a double inlet was to increase the volumetric capacity and mechanical efficiency by an appreciable amount. The effect of a reduction in the number of blades was a decrease in volumetric capacity and pressure developed, yet the mechanical efficiency increased somewhat with 48 blades. Figure 14 is for the same fan as that having 48 blades in Fig. 13, and shows what an evas6 (expanding) discharge on the fan will do in raising the resistance against which a given volume can be delivered or raising the volume delivered against a given resistance. The evas6 discharge connection increased in area with a taper to its sides of about 7 deg. until it was twice that of the fan outlet. For any given volume the brake horsepower is the same as for the fan of Fig. 13, the gain being in static pressure produced.
Figure 15 covers a steel-plate paddle-wheel fan with eight flat radial blades of relatively large radial depth and a single-inlet spiral casing. The side plates are inclined inwardly as in Fig. 20. The principal proportions are as follows: Number of blades, 8; radial depth of blades at intake, 0.145D; maximum radial depth of blades, 0.29D; axial length of blades (outer edges), 0.39D; axial length of blades (at intake), 0.46D; diameter at inlet of wheel, 0.71D; equation of spiral, R = r(l + 0.106a + 0.125); width of spiral, 0.5D; number of inlets, 1; diameter of inlet, 0.71D; area at cutoff point, 0.306D2; area at outlet, 0.306D2. While there is no!: much difference in mechanical efficiency between this fan and the multiblade fan of Fig. 12 there is a great difference in the pressures and volumetric capacities; much less space is required for the multiblade type when performing a given duty.